Construction machines, such as bulldozers and wheel loaders, generally have gear shifting mechanisms based on planetary gears. Through the gear shifting mechanism based on the planetary gears, switching between forward and reverse travel and speed changes are accomplished.
FIG. 10 shows an example of a planetary-gear type transmission (hereinafter referred to as "transmission") 10 which is mounted on a construction machine. The transmission 10 has a forward/reverse gear selector 20 and a speed gear selector 50 housed in a case 11. The forward/reverse gear selector 20 has a reverse sun gear 22 and a forward sun gear 24 which are fixed on an input shaft 13 connected to a driving source 12, such as a torque converter. A ring gear 26 is disposed around the reverse sun gear 22; a reverse planetary gear 30, which is supported by a reverse carrier 28, is meshed with the ring gear 26 and the reverse sun gear 22. The carrier 28 can be engaged with or disengaged from the case 11 through the reverse clutch (R clutch) 32.
The ring gear 26 is connected to a forward planetary gear 34 meshed with the forward sun gear 24 and a carrier 36 which supports the planetary gear of the speed selector 50 to be discussed later. A forward ring gear 38 meshed with the planetary gear 34 can be engaged with or disengaged from the case 11 through a forward clutch (F clutch) 40.
The speed selector 50 has a sun gear 52 for first gear, a sun gear 54 for second gear, and a sun gear 56 for third gear, which are fixed on an output shaft 14. A planetary gear 58, meshed with the sun gear 54 for second gear, and a planetary gear 60 for third gear, meshed with the sun gear 56 for third gear, are supported by the carrier 36 which supports the forward planetary gear 34. Further, the planetary gear 58 for second gear and the planetary gear 60 for the third gear are meshed with a ring gear 62 for second gear and a ring gear 64 for third gear, respectively, which are disposed outside thereof. These ring gears 62 and 64 can be engaged with or disengaged from the case 11 through a second gear clutch 66 and a third gear clutch 68, respectively.
A carrier 70, connected to the ring gear 62 for second gear, supports a planetary gear 72 for first gear. The planetary gear 72 meshes with the sun gear 52 for first gear and also with a ring gear 76 for first gear, which is disposed around the sun gear 52 and which is engaged with or disengaged from the case 11 through a clutch 74 for first gear. A reference numeral 15 denotes the inertia of a vehicle.
The transmission 10 having the construction described above engages the clutch 32 or 40 of the forward/reverse gear selector 20 to select forward or reverse travel of a vehicle and obtains a required reduction ratio by engaging one of the clutches 66, 68, and 74 of the speed selector 50. For example, to set for the forward second gear, the transmission 10 causes the forward clutch 40 to engage with the second gear clutch 66, fixing the forward ring gear 38 and the ring gear 62 for second gear on the case 11, and releasing the other clutches.
The following gives more detailed description of the operating procedure for shifting from the reverse first gear into the forward second gear.
In the case of the reverse first gear, the reverse clutch 32 engages along with the clutch 74 for first gear, and the reverse carrier 28 and the ring gear 76 for first gear are fixed to the case 11, the other clutches being released. Therefore, in the first step, the reverse clutch 32 and the clutch 74 for first gear, which fix the reverse carrier 28 and the ring gear 76 for first gear, respectively, to the case 11, are disengaged. In the second step, the hydraulic pressure to the clutch 66 for second gear and to the forward clutch 40 is gradually increased. The clutch torque capacity of the clutch 66 for second gear with respect to the transmission input torque is set at a larger value than the forward clutch 40; therefore, the clutch 66 for second gear engages earlier. At this time, the energy absorbed by the clutch 66 for second gear is a small value which is close to the inertia of the transmission 10. Then, the forward clutch 40, which engages later than the clutch 66 for second gear, absorbs a larger amount of energy to reverse the inertia of the vehicle. Conversely, the reverse clutch 32 reverses the inertia 15 of the vehicle and absorbs a larger amount of energy when a shift is made from forward gear to reverse gear.
Thus, the transmission 10 needs to absorb a large amount of energy to absorb the inertia 15 of the vehicle when a shift is made from forward to reverse travel or from reverse to forward travel. Accordingly, the area of contact of the clutches 32 and 40 of the forward/reverse gear selector 20 must be increased. Increasing the area of contact of the clutches 32 and 40, however, would pose such problems as those described below:
(a) The radii of the clutch discs of the clutches 32 and 40 must be increased, with a resultant larger transmission 10.
(b) A larger clutch disc means a larger loss horsepower during idling, resulting in poor fuel economy of an engine.
As a solution to the problems, there is a conceivable method wherein a steering brake is interlocked when the speed change involving the shift between forward and reverse is made, thereby reducing the load on the transmission clutches. In this case, the transmission torque in the clutch system of the transmission 10 will be as shown in FIG. 11, provided that the clutch of the speed selector 50 of the transmission 10 remains unchanged, wherein I.sub.1 denotes the input shaft equivalent inertial moment, I.sub.2 denotes the inertial 15 of a vehicle, i.e., the output equivalent inertial moment, .omega..sub.E (t) indicates the engine speed, and .omega..sub.C (t) indicates the output rotational speed of the transmission. Further, T.sub.E indicates the output torque of the engine, T.sub.C indicates the output torque of the transmission clutch, T.sub.B denotes the braking torque, and T.sub.L denotes the running resistance of the vehicle.
The motion of such a system can be expressed by equations (1) and (2) given below: EQU I.sub.1 (d.omega..sub.E /dt)=T.sub.E -T.sub.C ( 1) EQU I.sub.2 (d.omega..sub.C /dt)=T.sub.C -T.sub.L +T.sub.B ( 2)
Therefore, interlocking the brake at the time of shifting between forward and reverse causes the braking torque T.sub.B to be applied to output equivalent inertial moment I.sub.2. When the forward second gear is shifted to the reverse third gear, the quantities Q.sub.C and Q.sub.B of work from the friction between the transmission clutch and the brake clutch, until the shift between forward and reverse, can be given by expressions (3) and (4) shown below: ##EQU1##
Decreasing the calorific value of a transmission clutch means decreasing the value of Q.sub.C in expression (3). Accordingly, in order to decrease the value of Q.sub.C, the vehicle speed before making the shift is decreased so as to decrease the value of (.omega..sub.EO -.omega..sub.CO) of the first term of expression (3) or the value of the engine output torque T.sub.E is decreased to decrease the value of (T.sub.E -T.sub.C) of the second term, or the braking torque T.sub.B is increased to increase the value of (T.sub.C +T.sub.B -T.sub.L) of the third term.
The first term of expression (3) denotes, however, the status before the shift is made, and therefore it should not be a factor to be controlled. Decreasing the engine output torque T.sub.E of the second term causes a power line without the transmission 10 to often develop "engine stall", and therefore it should not be selected as a factor to be controlled. Hence, it is concluded that the shift between forward and reverse should desirably be made by interlocking the brake so as to apply braking torque T.sub.B of the third term and the transmission output torque, thus reducing the thermal load on the transmission clutches.
When, however, the brake is interlocked with the transmission 10, a failure to provide a correct interlock timing or to select a proper method results in other problems such as poor shift quality. Specifically, the output shaft torque T.sub.O, after braking is engaged, equals the sum of the clutch torque T.sub.C of the transmission 10 and the braking torque T.sub.B, provided the running resistance T.sub.L is zero. Hence, when a speed change, which involves the shift between forward and reverse, is carried out, the output shaft torque T.sub.O, after the braking is engaged when the brake is interlocked by an operator who depresses a brake pedal, changes as shown in FIG. 12 until a clutch of the transmission 10 completes engagement after a speed change command was issued.
Thus, the braking operation performed by the operator for making the shift between forward and reverse causes both the clutch torque T.sub.C and the braking torque T.sub.B to be applied at the same time, and these torques T.sub.C and T.sub.B linearly increase. This, in turn, causes the output shaft torque T.sub.O after braking to linearly increase. At a point where time t.sub.1 has elapsed since the braking was engaged, the vehicle speed reduces to zero and the running direction of the vehicle switches from forward to backward or from backward to forward, then the braking torque T.sub.B is not applied after that point and only the clutch torque T.sub.C of the transmission 10 is generated. Therefore, when the braking is discontinued, the output shaft torque T.sub.O reduces greatly, causing an uncomfortable shifting shock. Likewise, when making a full shift between forward and reverse, as in the case of making a shift from the forward third gear to the reverse third gear, even if the load is shared by the brake, the clutch oil pressure at time t.sub.2, at which the engagement of the transmission clutch is completed, remains high, resulting in a great difference .DELTA.T between the clutch torque T.sub.C and the passing torque at that point. As a result, there remains an unsolved problem in which the shifting shock takes place at time t.sub.2 when the engagement of the clutch is completed and a great peak torque is generated, leading to shortened service lives of gears and shafts.